Rotary impact tool



March l2, 1957 s. B. MAURER 2,784,625

` ROTARY IMPACT Toor.

Original Filed Jan. 5, 1951 5 Sheets-Sheet l t INVENTOR. SPENCER B. MAURER Flc. 3 FIG. 4 75M RNEY gi@ "gg ii W4 March 12, 1957 s. B. MAURER 2,784,625

ROTARY IMPACT TOOL Original Filed Jan. 5, 1951 3 Sheets-Sheet 2 F'IG. 9

IN VEN TOR.

SPENCER B. MAURER Q BY ATTORNEY March 12, 1957 s. B. MAURER ROTARYIMPACT Toor.

3 Sheets-Sheet 3 Original Filed Jap. 5, 1951 zum m .R RU mA fm MM @y V B. Nm w. R /fn N fw A E P S, f7, Y B om.. E n m: N:

ROTARY IMPACT TGOL Spencer B. Maurer, Highland Heights, Ghia Continuation of abandoned appiication Serial No. 204,674, January 5, 1951. This application March 25, 1952, Serial No. 278,438

13 Claims. (Cl. 81--52.3)

This invention relates to portable, power-operated, im pact wrenches and similar tools for driving bolts, nuts, screws, and the like or for applying a rotational impact force to other objects, all such tools being hereinafter generically referred to as impact wrenches.

This is a division of application Serial Number 204,- 674, filed on ianuary 5, 1951, for Rotary Impact Tool, now abandoned.

The usual tool of this character comprises three basic parts, namely, a prime mover, such as an air or electric motor; a rotatable hammer driven by the prime mover and having a high moment of inertia; and a rotatable output or work shaft, including an anvil member adapted to be struck by the hammer for imparting a high torque to the output shaft. The hammer member is disengageable from the anvil member to permit the prime mover to rotate the hammer member freely for imparting a high angular momentum thereto, and a clutch mechanism is provided for repeatedly engaging the rotating hammer and the stationary anvil to apply the momentum' of the hammer member to the anvil and to the output shaft. The clutch mechanism is also adapted to disengage the hammer from the anvil when the resistance to rotation of the output shaft has stalled the prime mover, thereby permitting free rotation of the hammer member for again building up its angular momentum. This cycle of operation is repeated until the power to the prime rover is cut off, the rapidity of the cycles and the magnitude of the torque applied being dependent upon a number of factors, including the rotational resistance of the object to be driven, the construction and mode of operation of the clutch mechanism, and the power and acceleration of the prime mover.

One of the principal problems involved in producing an impact wrench is to provide a driving connection between the prime mover and the impact hammer and at the same time protect the comparatively delicate parts of the prime mover from the relatively heavy shock loads due to rapid deceleration during impact.

The principal object of the present invention is to provide an improved rotary impact wrench wherein the power connection between the rotor and the hammer protects the motor against unduly high shock loads.

Another object of the present invention is to provide a power connection between the rotor and the hammer of a rotary impact wrench which reduces the stresses to which the rotor is subject is its operation, thereby permitting the use of much lighter rotor construction.

A further object of the present invention is to provide a much more reliable and rugged rotary impact wrench.

A further object of the present invention is to provide a rotary impact wrench which is less fatiguing to the operator due to lower vibrational forces being imparted to the operators hand.

A further object of the present invention is to provide a rotary impact wrench having very few parts, thereby achieving an inexpensive, yet reliable construction.

For a better understanding of the present invention,

arent together with other and further objects thereof, reference is had to the following description taken in connection with the accompanying drawings, and its scope will be pointed out in the appended claims.

in the drawings:

Fig. 1 is a cross-sectional view of a rotary impact tool the features of this invention;

Fig. 2 is a sectional View taken along line 2 2 of Fig. l, showing in particular the rotor construction;

Fig. 3 is a sectional view taken along line 3 3 of Pig. `l and showing in particular the air inlet ports to the rotor;

Fig. 4 is a sectional view taken along line 4 4 of Fig. l and showing in particular the reverse valve mechanism;

5 is a sectional view taken along line 5 5 of Fig. i

l showing the clutch control valve mechanism;

Fig. 6 is a sectional view taken along line 6 6 of `Fig. l showing the hammer;

Fig. 7 is a sectional view taken along line 7 7 of Fig. l showing the anvil;

Figs, 8 and 9 are sectional views of a modified form of the invention;

Fig. l0 is a cross-sectional view of a rotary impact tool embodying an adjustable torque control mechanism;

Fig. l1 is a cross-sectional View taken along line 2 2 of Fig. l0;

Fig. l2 is a cross-sectional View taken along line .3 3 of Pig. ii; and

Fig. 13 is an isometric view of a small portion of the impact tool.

in the past, many impact tools have had very noticeable rotary vibration or twist arising from the reaction torque of the motor. This twist or torque reaction is imparted to the operators hand and is often of sufficient magnitude to make it ditlicult for the operator to control the tool. Such a condition makes it very fatiguing to operate an impact tool causing lowered tool and operator output.

impact tools have been made using rather delicate motor parts; the rotor and vanes being made as light as possible in order that the Huid pressure will accelerate them as rapidly as possible.

High torque reaction and vibration are very hard on the delicate parts of an impact tool, particularly on the vanos. The rotor in some tools actually rebounds after each impact blow and turns backwards relative to the housing against the forwardly acting pressure of the air in the motor, When this rebound occurs, the reaction force on the operator is considerably higher than the maximum torque reaction of the motor when allowed to run in a forward direction. Therefore, the forces which produce the most objectionable` handle reaction are those set up by the rebound action of the rotor. The output torque of the motor in a forward direction is the theoretical twisting force produced by the fluid pressure minus the frictional resistance of the various parts of the motor, principally the sliding vanes rubbing against the core of the motor. When Linder rebound condition the rotor is forced to rotate backwards against the forwardly acting fluid force, the backward reaction on the housing equals the fluid pressure torque plus the resisting torque caused by the frictional forces between the vanes and motor bore. inasmuch as the frictional forces referred to above may be nearly as large as the forwardly acting iiuid pressure forces, it is easy to see how the handle reaction, when the rotor is rebounding, may be several times as great as the handle reaction when the rotor is turning in a forward direction under the influence of the fluid forces of the motor.

The frictional forces between the blades and the rotor are much greater during this rebound period than during the normal forward motion. rl`he sliding action of the blades in the rotor is impeded by the unfavorable direction of motion of the rotor in relation to the motor bore and the applied fluid pressure. rl`his leads to greatly increased stresses on the blades and increased wear on the mating parts.

By placing a friction clutch between the motor and the hammer unit which will slip at a torque slightly greater than the normal output torque of the motor, the tendency of the rotor to rebound is eliminated because the rotor continues to rotate in a forward or ctie-n willie the hammer is delivering the impact blow. The hammer is free to rebound after delivering the impact blow while the rotor ispstill going in a forward direction; and the kinetic energy of the rotor is dissipated in the form of heat in the friction clutch` Even though the frictional torque at which the clutch slips is of a greater value than the objectionable torque reaction that we are trying to eliminate, the frictional slip clutch can eliminate this objectionable rebound torque reaction. The reason for this is that the kinetic energy of the rotor at the time of impact is suiciently large to insure a continued forn ward rotation of the rotor as it is decelerated by braking action of the slipping friction clutch as the hammer is rapidly brought to a standstill. The hammer may even rebound while the rotor is moving forward, the slipping of the friction clutch permitting this condition.

Three types of friction clutches capable of greatly reducing the torque reaction of an impact tool are shown and described, one on each sheet of drawing.

The friction clutches which are shown on sheets 2 and 3 of` the drawings preferably are made of tool steel, carburized and hardened steel, sintered metals such as frictional clutch materials, rubber, leather or the like; and the frictional clutch which is shown in Figure l preferably is made of rubber, leather or the like.

In the drawing there is shown a rotary impact wrench embodying an improved air motor and a new and irnproved clutch mechanism, a portion of the fluid under pressure (in this case compressed air) normally supplied to the tool for driving the motor being bypassed to the clutch mechanism, where it is employed to eect both engagement and disengagement of the hammer and anvil. Centrifugal force is employed to assist in the control of the clutch, by solely operating a simple valve in the compressed air system. The primary forces responsible for an operation of the clutch are applied to the mechanism entirely by compressed air. As shown in Fig. l, the tool is mounted in a housing that may conveniently comprise an integral hammer and motor housing section l@ and a separate housing section il' containing or mounting the parts of the device driven by the motor. The two housing sections w and il may be suitably threaded together or otherwise secured in any conventional manner.

The motor, the controls therefor', and the compressed air supply line leading to the motor may, for the most part, be of any conventional design and need not be described in detail. Essentially this portion of the device includes a coupling l?, for a compressed air hose 13 that leads into the handle portion of the housing section' l@ and to a main control valve contained therein (not shown). A trigger td is mounted on the handle portion for actuating the main control valve, from which a main air conduit i6 leads toward the motor. The air in the'main air conduit lo passes through a reverse valve 17 interposed between the main air conduit le and the motor for controlling in a well-known manner the supply of air selectively either to the manifold its for forward operation of the motor or to the manifold i9 for reverse or clockwise rotation of the motor. rhe air enters the cylindrical air chamber 2l either through the ports 22 or the ports 2li, drives the rotor by pressure on the vanes V24, and is exhausted through the exhaust port 25 and exhaust conduit Z7'. rlfhe rotor 23 is eccentricallgr mounted with respect to the liner C in the rotor chain.--

' about its axis.

ber, and the varies Z4 are slidably mounted in the radial slots in the rotor and are urged outwardly against the liner of the rotor chamber by air pressure through ports A for forward motion and through ports B for reverse motion. The air entering the rotor chamber between the adjacent pairs of vanes moves the vanes about the rotor axis in the direction to enlarge the space therebetween permit expansion of the air as it travels toward the exhaust port 2e. The valve i7 that controls the direction of the rotation of the rotor is manually cont de valve 29 presently shown in position forward rotation.

The rotor is mounted on a hollow drive shaft 3l by means of a rubber coupling sleeve 32 which frictionally engages both the rotor 28 andthe drive shaft Eiit. "ifo assemble the rotor and the drive shaft, the rubber sleeve 32 is slipped on over the drive shaft 3l with one end of the rubber sleeve against the shoulder D on the drive shaft and with the other end of the rubber sleeve 3&2 extending beyond the location of the end of the rotor The rotor 28 is then slide into position over the rubber sleeve 32 and a collar 33, which fits tightly around the metal drive shaft 3ft, is slid into position against the end of the rubber sleeve 32 to compress the rubber and cause the rubber to frictionally engage both the drive shaft 3l and the rotor 2S to provide a frictional coupling between the two members. Ball bearings 3d are provided on the forward and rearward ends of the drive shaft and a snap ring E is connected to the collar 33 to hold the rearward ball bearing 34 in place. This provides a driving connection of sufficient torque to transmit 'the normal motor torque to the hammer with essentially no slippage but to allow slippage during periods of rapid deceleration of the hammer and drive shaft so that the rotor body is never rapidly decelerating and the kinetic energy and momentum of the rotor body are dissipated in the form of heat at the slipping surface, the advantage of this being that the rather light and delicate rotor body is not subjected to hezvy shock loads or stresses.

The output shaft 37 is journalled in a bushing 33 carried by the housing section lill and terminates at one end in a suitable work driving tip 39. The bushing 3S is sufficiently long to hold the output shaft 3'7 firmly in alignment with the axis of the drive shaft 3l and to permit the clutch assembly to be supported solely by and between the output shaft 37 and the hollow drive shaft Si, in a manner hereinafter described. The output shaft 3'7 is provided with a circumferential shoulder portion il for retaining the shaft in the bushing 38. A projection 43 from the shoulder' portion il of the output shaft 3'7 constitutes an anvil and is shaped on opposite sides to provide identical impact surfaces 44 against which a hammer may strike to drive the output shaft in either direction The hammer d8 includes a projection 4S which may project into the path of radial projection 43 of the anvil member. The clutch assembly is constructed around a relatively massive hollow hammer 48. The forward end of the hammer body dit is splined to the spindle i9 by spline teeth Su and is supported for rotary motion with the spindle and is capable of axial V.motion relative to this spindle. A steel ring 5l is mounted between the hammer d@ and the spindle i9 at the rearward end thereof and a snap ring 52 connects the steel ring 5l to the hammer Vbody iti so that the hammer body 48, the steel ring Si, and the snap ring 52 lmove 'axially as a unit with respect to the spindle 49. An annular opening 5s is provided between the hammer d8 and the spindle 49. Air under pressure fills this opening 53 and exerts a driving force directed forwardly against the face 5d of the hammer Another annular opening 55 is provided between the hammer 4S, the steel ring 5i, and the spindle 59 When air under pressure is admitted to the opening it drives the hammer unit comprised of the hammer dis, the steel ring 5i, and the connecting ring 52 toward the Een rear of the tool to disengage the projecting lug 48 from the anvil face 44.

A valve mechanism 60 is provided for lcontrolling the air into the annular space 55 to intermittently connect the space 55 to the source 16 of air under pressure and to atmospheric air to permit the air under pressure in the annular space 53 to slide the hammer 48 forward. Air under pressure is provided to the annular space 55 from the source 16 through the valve 17, through the drilled passageway 61 to the hollow drive shaft 31 to the .space 62 under the valve body 6i). From the space 62 this air under pressure passes through drilled passageway `63 to the annular space 53. The air in annular space 53 fills the space 64 between the hammer shell 48 and the top surface of the valve 6d, thereby balancing the air pressure acting on both sides of the valve 60. The valve '60 includes a stem portion 65 which extends into a passageway 66 in the body of the spindle 49. The passageway 66 extends to an opening 67 which is vented through `an annular space 68 and drilled passageways 69 and 70 and out through the front of the output shaft 39. While the air pressures on the top and bottom surfaces of the valve 6G are balanced, the areas on the top and bottom surface over which their air pressure is active is slightly greater in favor of the outer surface due to the stem 65 blocking off a portion of the air under pressure in space 62. The net result is a slight force biasing the valve 60 inwardly toward the axis of the tool.

Fig. 1 shows the valve 6d in its outward position where it is forced by centrifugal action which overcomes the slight air bias. In the position shown in Fig. l with the valve 60 in its outward position, the annular space 55 is vented to atmosphere through the opening 70 through annular opening 7i around the valve 60 and the opening 72 in the spindle body 49 to the vented space 68. This causes the air under pressure in the annular space 53 to push the hammer 48 forward into engagement with the anvil 43 thereby using the inertia of the hammer for striking a blow for driving or removing nuts, bolts, or the like. At the termination of an impact, the hammer and `spindle have lost their angular velocity and are standing essentially still. Therefore, the centrifugal force previously acting on valve 60 has disappeared and the air bias becomes predominant and moves the valve body 60 radially inward to a position where port 73 is nowfconnected to the live air of annular chamber 53 and 64. This allows air under pressure to reach the larger areas of chamber S which produces a rearwardly acting force on the hammer body 4S which exceeds the opposing force from the air in chamber 53, thus causing the hammer body to be moved in a rearward direction and disengages the impact projection 48 from the anvil 43 thus allowing the motor to again accelerate the hammer assembly.

The clutch is lubricated periodically by grease supplied through grease fitting 8i), through several holes Si drilled in the bushing 3S, and through a hole SZ in the output shaft to the proximity of the impact jaws. In this manner grease is supplied to all of the bearing surfaces in the front end of the impact clutch. in order to retain the necessary minimum amount of grease in the clutch chamber at all times, the vent 69 in the forward end of the spindle is the only opening through which grease could leak to the outside of the tool; but since this opening is essentially centrally located within the clutch chamber, it is impossible for all of the grease to run out or to be blown out of the clutch housing. As the tool is tilted to the various operating positions, there is always a suflicient volume of space in the clutch housing which is at a lower level than the opening of this passage 69 and therefore the grease will not tend to run out.

Figs. 8 and 9 show an alternate embodiment of the friction drive feature of the invention and an alternate clutch mechanism for coupling the motor to the hammer.

' The motor, the controls therefor, and the compressed vair `supply line leading to the motor may, for the most part,

CIK

be of any conventional design and are not here described in detail. For a brief description of the parts, reference may be made to the device shown in Figs. l to 7.

in the embodiment shown in Figs. 8 and 9, the rotor 28 and the drive shaft 3l are connected by spline 30 to a cone driver 35 for causing the cone driver to rotate with the motor 28 and drive shaft 31. The hammer 48 has a conical surface portion 36 which is adapted to be frictionally engaged by the inclined face of the cone driver 35. A spring 40 biases the cone driver into engagement with the conical surface portion 36 of the hammer 48. The spring 40 is held in place by a retaining ring 42 the outer edge of which is locked in a groove 45 in the hammer 48. Rotation of the cone driver 35 causes the hammer 48 to rotate.

The hammer 48 is comprised of a clutch body 85, a piston 86, a valve pin 87, a centrifugal weight 88, and an impact pin 89. The clutch body is mounted for rotary motion by means of the cone driver at one end and by means of the output shaft 9G at the other end. The clutch body S5 has a hollow central portion and has a groove 91 which receives a portion of the Icentrifugal weight 88 to permit the weight to slide radially across the tool. One end of the valve pin 87 is in engagement with an upstanding portion 88' of the centrifugal weight and the other end of the valve pin is mounted for sliding motion within a bore 92 in the piston 86. The valve pin 37 includes an enlarged shoulder portion 93 which sealingly engages the walls of the bore 92. The piston 86 is mounted within the hollow central portion of the clutch body 5S :and is connected to the impact pin 89 by means of the projecting lug 94 which fits in a slot 95 in the impact pin. The piston 86 and the impart pin 89 are mounted for axial motion with respect to the tool. When the impact pin is in the forward position shown in the drawing, its front end is in engagement with the impact receiving portion 96 of the output shaft 90, Ias described in detail in the other embodiment of the invention.

When the impact pin is in its backward or withdrawn position, its front end is disengaged from the impact receiving portion 96.

The operating cycle is as follows: starting from the position shown in the drawing where the torque resistance of the work is greater than the output torque of the motor and the motor is stalled. The impact pin is in engagement with the output shaft. Air pressure in the hollow recess 97 in the piston 86 reaches the shoulder portion 93 of the pin 87 through port 98 and constantly urges the ,pin 37 against the portion 88 of the centrifugal weight 83 to force the weight radially to the right. This moves the shoulder portion 93 past the port 99 to allow compressed air from port 98 to travel through port 99 to the chamber Mill at the front face of the piston S6. This pushes the piston 36 rearwardly, carrying with it impact pin 69, thus disengaging the impact pin from the output shaft 96 and allowing the rotor 2S to accelerate the hammer unit 48. Before the hammer unit. rotates one complete revolution, the centrifugal force acting on the weight 86 becomes sufficient to overcome the fluid force on the shoulder 93 of the valve pin S7 causing the pin 87 to move radially inwardly to a position such that the shoulder portion 93 effects a seal between the ports 98 and 99 thus disconnecting port 99 from its source of fluid pressure and opening port 99 to atmosphere pressure within the clutch housing thus allowing the vair pressure in chamber i0@ previously acting on the clutch face to disappear. The air pressure in chamber 97 causes the piston to move forwardly forcing the impact pin into engagement with the output shaft 90, so that by the time the hammer unit has completed one revolution the impact pin S9 is in position to engage the impact receiving face 96 of the output shaft 96.

The torque required to produce slippage between the cone driver 35 and the conical surface portion 36 of the hammer 48 is greater than the maximum output torque 7 of the rotor, but is not great enough to prevent slippage during an impact due to the momentum of the rotor rihis slippage prevents excessive impact shock loads from reachingthe rotor 2S.

The tool is assembled in a twcpart threaderbtoeether housing which is locked togefher by a pinch-"c The housing section 1116 encloses the motor unit and the housing section 1117 encloses the clutch parts. The two housing sections threaded together'. One of them, for example housing section 1116, carries the female threads, and is provided at tie thread area with spr 'i outwardly estending lugs 1113, 11,29 with a s nl running up through the threads. A threaded bore 19d extends through the lugs 1123, 1113* in a direction perpendicular to the axis of the tool and preferably at a location below the threads in the housing portions Screwing the pinch boit tightly into posi housing portions have been threaded to. the lugs 1nd, 1139 togethe greatly reasing the radial force between tie two threaded l'iusing portions, setting up a higher frictional locking force than is otherwise obtained.

With reference to l) of the drawing, the rotary impact tool comprises a housing 11i) which, for convenience in assembly, may be made in several serI 'ons and suitably connected together after assembly of ine tool within the housing.

An anvil 111 is rotatably mounted within the housing 1111 and is journalled at 1"?. for rotary motion with respcct to the housing. An end of the anvil 111 is square and protrudes outside of the housing and is provided with a locking pin 11Ei for connection to sockets which are well kno-wn in the art. rl`he other end of the anvil 111 is provided with an impact Vreceiving surface LT1/1 which is adapted to be engaged by impact delivering surface comprising a portion of a massive hammer element 116. The hammer 115 is comprised of a cylindrical member 117 integrally connected to the impact delivering surface 115 and adapted for rotation with a central spindle 1 and a cross valve member 119 which is carried by the spindle. Cross valve 119 is slidably mounted in an is constrained to rotate in unison with the spindle and has limited travel in a radial direction relative to the spindle. As shown in Fig. l2, the lug 123 on the interior surface of cylinder 117 provides a driving connection with the spindle 118. The spindle 1125 has hole 12.41 bored through it in a radial direction in which is slidably positioned the cross valve 119, The cross valve 119 is composed of a cylindrical valve portion 125 and an eccentric weight 126 on one end, As is shown in Fig. ll, the cross valve member 119 has a clearance hole 121i extend ing through it, and through the hole 1215 there entends a portion 121 of massive means 122. The massive means 122 is inserted in a central bore 129 at the rear ond of spindle 118 and is mounted for rotation with, or relative to, the spindle 11S by means ofthe balls 127 and cooperating cam grooves 12S and 128 in opposed surfaces of the spindle 11S andthe massive means 122.

The front end of spindle 11S is journalled in bere 1319 of anvil 111 and is supported at its rear end by the cone driver 131 which in turn is supported on a protruding portion of the rotor shaft 13:2. The driving connection between the rotor shaft and the cone driver comprises interlocking spline teeth 133, and the driving connection between the cone driver 131 and the spindle 118 conprises a conical frictions] surface 134. Closure plate s is positioned within the rear end of the cylinder and is held in place by retaining ring 1296. A seal ring 137 effects a seal between the closure plate and the cone driver' 131 and prevents loss of compressed air. rthe closure plate 135 rotates and moves asiairy with the cylinder 117. Rotor shaft 132 is journalled in ball bearings 14E-l) and 1111 mounted in the fro-nt end plate 142 and rear end plate 143 of the motor 14st.

The motor is of the usual sliding vane and eccentric cylinder construction and is driven by compressed air supplied from the throttle 1511 through passageway 151 to a reversing cap 152 which supplies live air through annular passageway 153 to ports at the rear end plate 143 for leither forward or reverse operation. Reversing cap 152 also contains a forward-facing internal bore 155 in which is mounted a portion of air tube 156. The air tube 156 extends from the bore 155 into one end of the hollow rotor shaft 132. An adjusting screw 157 is threaded into a central hole through air tube 156 and nous on the iorward end against spring button 158 which in turn is seated against spring 159. Spring 159 is mounted within the hollow rotor shaft and its forward end abuts against a rearwardly projecting portion of the massive means 122.

A portion of the live air which is supplied to the motor from the annular passageway 153 of the reversing cap 152 is conducted through passageways 165 and 166 to chamber 167 at the rearward end of the air tube 156, thence down through a central hole 168 in the air tube 156 and through the hollow adjusting screw 157 radially outward through passageway 169 to the annular space 171D between the rotor shaft and the spring button 158 and again radially downward to a central hole 171 in the spring button and forward through and around the spring 159 to a central hole 172 in the massive means 122, then through a communicating hole 176 in piston 175 into chamber 177 located in the forward portion of spindle 118.

As shown in Fig. 1l, chamber 177 communicates through passageway 173 with an annular space 179 between hammer cylinder 117 and spindle 118. As long as the throttle 15@ is open, live air is supp-lied to this chamber 179 and is conned within this chamber by seals and 151. Fluid pressure within chamber 179 exerts a forwardly acting force on the hammer cylinder 117 and a rearwardly acting force on the spindle 118. The spindle 118, however, is not allowed to move in an axial direction while the hammer cylinder 117 is periodically moved forward and back in an axial direction to obtain the intermittent engagement and disengagement of the impact surfaces. Fluid pressure in chamber 179 also eX- erts a force on each end of cross valve 119. Since the diameter of the bore 124 at the head end of the cross valve is larger by several thousandths of an inch than the diameter at the opposite end of the cross valve, there is an 4inherent biasing force on the cross valve 119 tending to hold it in a retracted position so that the head 126 is radially inward toward the axis of the tool. Clearance hole 1211 is vented to atmospheric pressure through :txal holes 183 in massive means 122 and back through the grooves between the spline teeth 133 on rotor shaft 132 to annular chamber 184 in the front end plate 142, and from there radially outward through drilled passageway 185 to main exhaust passage 186. Groove 196, Fig. l2, on the small end of cross valve 119 forms a connecting passageway between the annular space 179 and drilled port 191 thus allowing live air to enter passage way 191 and to travel through angular hole 192 to a rear annular chamber 193 between the spindle 118 and the closure plate 135. Fluid pressure in this chamber, 193, exerts a rearwardly acting force on the hammer cylinder 117 and a forwardly acting force on the spindle 118. Since the spindle is restrained from axial motion and also since the effective area of the closure plate is greater than the effective area of the forward annular chamber 179, the rearwardly acting forces predominate and cause the hammer cylinder 117 to move in a rearward direction, thus causing disengagement of the impact delivering surface 115 from the impact receiving surface 114. This allows the motor 144 to accelerate, in a rotary direction, the massive hammer element 116. Eccentric head 126 of cross valve 119 soon attains sufficient centrifugal force to overcome the air pressure bias which is acting on valve 119, causing the valve to move radially t 9 outward to the position shown in Figs. 11 and 12. In this position, the groove 190 on the small end of the valve connects passageway 191 with the vented chamber 195 in the center part of the spindle 118. Thus the supply of live air to the rearwardly acting annular chamber 193 is cut off and this chamber is vented to the atmosphere through the angular hole 192 and the drilled port 191 thus dissipating any forces previously present on closure plate 135. This allows the engaging forces of the annular space 179 to again predominate and cause the hammer cylinder 11'7 to move forward so that the impact delivering surface 115 is again positioned to engage impact receiving surface 114 when the proper amount of rotation has occurred. When impact occurs between the hammer and the anvil, the hammer element 116 as a unit is decelerated by the resisting torque of the anvil as it tends to turn the nut or bolt that is being driven.

Spring 159, acting through massive means 122, holds the piston 175 in engagement with valve seat 197 of the spindle. This confines the live air in chamber 177 and prevents this air from entering the annular chamber 199 between the piston 175 and the spindle 118. There is suiiicient clearance 199 between the O. D. of piston 175 and the I. D. bore of spindle 118 to allow a small rate of leakage from chamber 198.

While the tool is operating on a nut or bolt which turns readily, the massive means 122 is coupled to the hammer 116 by means of the force of the spring 159 acting through balls 127 and inclined grooves 128, 128. When the nut or bolt does not turn readily, causing the tool to tend to exceed a set maximum torque output, the massive means becomes uncoupled from the hammer as the inertia of the massive means overcomes the torsional force of the spring exerted on the massive means through the balls 127 and cam grooves 128, 128. When the massive means overrides the hammer, the ba'lls 127 roll between the opposing cam grooves 123, 128' causing the massive means to move axially in a rearward direction along a helical path parallel to the path of the cam grooves. The movement of the massive means 122 in the rearward direction causes the seating force to be removed from the valve seat 197 and allows air from the chamber 177 to enter into chamber 198 at a rate great enough to exceed the leakage from this chamber through clearance 199, thus applying an overbalancing fluid pressure force against the forward end of piston 175 which forces the massive means 122 to move rearwardly against spring 159 until the rear portion of piston 175 contacts the side of sliding cross valve 119. This prevents any further compression of spring 159 and sets up a high frictional or binding force on the cross valve 119, thus preventing the normal air biasing force from retracting said valve. This prevents disengagement of the impact surfaces and maintains the unit in a stalled condition until throttle valve 150 is manually closed.

When throttle valve 150 s manually closed, all air beyond the throttle valve is dissipated through previously mentioned exhaust channels and clearances, thus allowing the spring 159 to return the massive means 122 and the piston 175 to their original position so that a seal is again established in the valve surface 197. When the throttle is again opened, the tool will again operate in a normal fashion producing repeated impact blows until the force of said blow and, consequently, the rate of deceleration of the hammer again becomes suiiicient to cause the massive means 122 to rotate forwardly ahead of the spindle, again causing the control mechanism to operate.

The torque at which the control mechanism operates can be varied by varying the initial force of spring 159. A slot 200 in the rear end of adjusting screw 157 can be reached from the rear of the tool by removing plug 201, and a screwdriver inserted in said slot is used to turn the screw either forwardly in to increase the spring pres- Cil 1G r sure or in the opposite direction to reduce the spring pressure.

The driving connection between the shaft 132 0f the driving motor and the hammer assembly 116 is through a conical friction clutch at surface 134. The value of the frictional torque at which this clutch slips has two diiferent levels during one cycle of operation. When the hammer 117 is in its forward position ready to deliver an impact blow, as shown in Fig. 10, the clutch loading force is set up by the air pressure force of chamber 179 acting on surface 101 and holding the hammer in its forward position. The forward motion of the hammer is limited by the closure plate abutting against the rear face of the driving cone 131. When the centrifugal valve 119 is retracted and live air pressure is admitted to the rear disengaging chamber 193 through conduits 191, 192 closure plate 135 is forced backwards away from cone 131 and the cone loading force becomes the direct pressure force acting on the area of cone 131 exposed to chamber 193. For this reason the slip torque of the friction driving clutch is at a relatively low value when the impact blow is delivered and at a higher value when disengagement occurs and the hammer element is to be accelerated in a rotary direction. to continue to rotate while the hammer is delivering the: impact blow and the residual kinetic energy of the rotorl when disengagement of the hammer occurs is available: to help accelerate the hammer again.

While there have been described what are at present considered to be the preferred embodiments of this invention, it will be obvious to those skilled in the art that: various changes and modifications may be made therein without departing from the invention, an-d it is, therefore,l aimed in the appended claims to cover :all such changes and modifications as fall within the true spirit and scope of the invention.

What is claimed is:

l. In a rotary impact tool; a rotatable output shaft; rotatable hammer means including rotatable massive means and releasable clutch means located between said massive means and said shaft, said clutch means including hammer and anvil surfaces arranged to engage each other to deliver an impact blow, and automatic means for disengaging said hammer and anvil upon termination of said blow; rotatable motor means having a given mass, means for accelerating said motor means, frictional driving connection means effectively coupled between said motor means and said rotatable hammer means for coupling said motor means to said hammer means, said frictional connection means including a pair of mating frictional surfaces capable of transmitting accelerating forces from the motor to the hammer with substantially no relative slippage and capable of unidirectional slippage with respect to each other due to the inertia of the motor mass when the hammer strikes the anvil to a new different relative position which is maintained until a subsequent blow of the hammer means against the anvil.

2. A rotary impact tool as set forth in claim l, further characterized by said frictional driving connection means including a drive shaft and frictional means, said frictional means being in frictional engagement with said motor means and in frictional engagement with said drive shaft.

3. A rotary impact tool as set forth in claim l, further characterized by said frictional driving connection means including a drive shaft and by said frictional means being in engagement with said drive shaft and said rotatable hammer.

4. A rotary impact tool as set forth in claim 3, further characterized by means rigidly connecting said frictional means to said shaft, and by said frictional means being in frictional engagement with said rotatable hammer.

5. A rotary impact tool as set forth in. claim 4, further characterized by uid pressure means exerting a force on This allows the rotor' said frictional means urging said frictional means into frictional engagement with said rotatable hammer.

6. `A rotary impact tool as set forth in claim 5, further characterized by iirst fluid pressure means exerting a force on said frictional means at one time urging said frictional means into frictional engagement with said rotatable hammer, and by second uid pressure means exerting a force on said frictional means at another time urging said frictional means into frictional engagement with said rotatable hammer.

7. A rotary impact tool as set forth in claim 6, further characterized by clutch means causing said iirst fluid pressure means to exert said force on said frictional means when said hammer and anvil surfaces are in engagement, and for causing said second iiuid pressure means to exert said force on said rictional means when said hammer and anvil surfaces are out ot engagement.

8. A rotary impact tool as set forth in claim 7, further characterized by said second fluid pressure means exerting a greater force on said frictional means than said rst fluid pressure means.

9. in a rotary impact tool; a rotatable output shaft including an anvil; rotatable hammer means; a Clutch in engagement with said hammer means to periodically cause said hammer means to impart a blow to said anvil; automatic means for disengaging said hammer means from said anvil upon termination of said blow; a rotatable motor; and rictional driving means effectively coupled to said motor and effectively coupled to said hammer means for frictic-nally transmitting without substantial slippage rotational acceleration forces from said motor to .said hammer means, said frictional driving means slipping when said hammer means decelerates rapidly during impact.

10. A rotary impact tool as set forth in claim 9, further characterized by said rictional driving means comprising two frictionally engaging members, and by resilient means exerting a force against onerof said members urging it into frictional engagement with the other said member.

11. A rotary impact tool as set forth in claim 10, fur- 12 ther characterized by said resilient means comprising fluid pressure means. i

12. In a rotary impact tool; a rotatable output shaft; rotatable hammer means including rotatable( massive means and releasable clutch means located between said massive means and said shaft, said clutch means including hammer and anvil surfaces arranged to engage each other to prevent relative rotation between said hammer and anvil surfaces when said hammer and anvil surfaces are in engagement, and automatic means for disengaging said hammer and anvil upon termination of an irnpact; a rotatable motor for driving said hammer means; frictional driving connection means comprising two frictionally engaging members one of which is in driving engagement with said motor and the other of which is in driving engagement with said rotatable hammer means for frictionally transmitting rotational forces from said motor to said hammer means, and coupling means exerting a coupling force against one of said engaging members urging it into frictional engagement with the other said engaging member, the value of said coupling force being such as to substantially prevent slippage between said two frictionally engaging members during acceleration of the hammer by said motor and permitting slippage under the greater torsional forces encountered as said motor decelerates during impact.

13. A rotary impact tool as set forth in claim 12, further characterized by said coupling means comprising resilient coupling means.

References Cited inthe file of this patent UNlTED STATES ?ATENTS 2,049,273 Pott July 28, 1936 2,219,865 Fitch Oct. 29, 1940 2,219,883 Amtsberg 2 Oct. 29, 1940 2,261,204 Amtsberg Nov. 4, 1941 2,476,632 Shai July 19, 1949 2,533,703 Wilhide et al Dec. 12, 1950 2,576,851 Newman Nov. 27, 1951 2,583,147 Kaplan Jan. 22, 1952 

